The present invention relates to a two-stroke cycle gasoline engine, and, more particularly, to a two-stroke cycle gasoline engine adapted for use with automobiles.
A two-stroke cycle engine has theoretically the advantage that an engine of a certain size can generate a greater power than a four-stroke cycle engine of a bigger size because the two-stroke cycle engine has twice as many work cycles per revolution as the four stroke cycle engine. In fact, however, a conventional two-stroke cycle gasoline engine employing a carburetor has drawbacks, such as: that it has high fuel consumption as compared with a four stroke cycle engine due to the loss of fuel-air mixture caused by the direct escape, i.e. blow-out, of scavenging mixture to the exhaust manifold during scavenging; and that it cannot generate such a high power as expected from the fact that it has twice as many work strokes as the corresponding four-stroke cycle engine, due to the fact that the scavenging is still insufficient.
As methods of scavenging in two-stroke cycle engines are conventionally known cross scavenging, loop scavenging, and uniflow scavenging. In this connection, if the amount of scavenging mixture is increased so as to improve scavenging efficiency, uniflow scavenging is considered to be most desirable, in order to obtain the highest scavenging efficiency without causing direct escape of the scavenging mixture to the exhaust manifold. In view of this, and in view of the aforementioned drawbacks, the actual application of two-stroke cycle gasoline engines has been conventionally limited to the field of small-size engines in which simplicity of structure and low manufacturing cost are essential conditions. Therefore, conventional two-stroke cycle gasoline engines presently used generally employ crankcase compression for scavenging. However, scavenging by crankcase compression cannot deliver a sufficient amount of scavenging mixture, thereby causing incomplete scavenging, which leads to a relatively low volumetric efficiency.
In view of the fact that such a low volumetric efficiency is the principal cause of the poor output power of conventional two-stroke cycle gasoline engines, in a preceding patent application Ser. No. 917,244 we have proposed a two-stroke cycle gasoline engine particularly suitable for use as an automobile engine, which comprises at least one two-stroke cycle power cylinder--piston assembly incorporating uniflow scavenging and two horizontally opposed pistons, and at least one scavenging pump cylinder--piston assembly of the reciprocating type, with or without incorporating crankcase compression, wherein the total stroke volume of the scavenging pump means is 1.35 to 1.85 times as large as that of the power cylinder--piston assembly, so that the volumetric efficiency is substantially increased so as to generate high power output when compared with conventional two-stroke cycle gasoline engines.
Furthermore, in view of the fact that, even when a separate pump cylinder--piston assembly is employed as proposed in the abovementioned former application, if the conventional crankcase compression is also incorporated, the operational phase relation between the power cylinder--piston assembly and the scavenging pump means is substantially restricted, we have proposed, in our former application Ser. No. 917,241, not to utilize at all crankcase compression, and to provide a two-stroke cycle gasoline engine which comprises at least one two-stroke cycle power cylinder--piston assembly of the reciprocating type and driven by said power cylinder--piston assembly in synchronization therewith with a phase difference therebetween, wherein the total stroke volume of said pump cylinder--piston assembly is between 1.15 and 1.65 times as large as that of said power cylinder--piston assembly, and said phase difference between said power and said pump cylinder--piston assemblies is so determined that the top dead center of a pump cylinder--piston assembly is, as viewed in the crank angle diagram, in a range between 15.degree. in advance of and 15.degree. behind the midpoint between the bottom dead center and the scavenging port closing phase point of the power cylinder--piston assembly to which it supplies scavenging mixture, thereby substantially improving its scavenging efficiency when compared with conventional two-stroke cycle gasoline engines, so that the engine can generate high output power and is suitable for use as an automobile engine.
In either of the abovementioned formerly proposed two-stroke cycle gasoline engines, in order to improve scavenging efficiency, the amount of scavenging mixture is increased by employing an additional or separate pump when compared with the conventional scavenging dependent only upon crankcase compression, so that the volumetric efficiency of the power cylinder becomes as high as 75%-100%. In this connection, in the case of a power cylinder--piston assembly having two horizontally opposed pistons and incorporating uniflow scavenging, when the path of the flow of scavenging mixture from the scavenging port to the exhaust port, and its length, which are determined by the geometry of the power cylinder and the scavenging and exhaust ports provided at its scavenging and exhaust sides, respectively, and the manner of flow of the scavenging mixture in the power cylinder which is ejected into the power cylinder through the scavenging ports and generally flows along a helical path, are given, the speed of scavenging, i.e. the speed at which the scavenging mixture moves from the scavenging port to the exhaust port while urging the exhaust gases remaining in the power cylinder, is determined by the difference between the pressure of the scavenging mixture and that of the remaining exhaust gases, and by the amount of scavenging mixture which backs up the scavenging mixture introduced into the power cylinder, i.e. by the total amount of scavenging mixture, and has no direct relation to the rotational speed of the engine. On the other hand, however, the time which lapses from the time point when the scavenging port is opened to the time point when the exhaust port is closed becomes shorter as the rotational speed of the engine becomes higher. Therefore, when so-called engine matching is so performed that, when the engine is operating at that load and rotational speed which are most frequently employed, the exhaust port is closed when the scavenging mixture has just pushed the exhaust gases out of the exhaust port (such a rotational speed is called "matching rotational speed"), assuming that the volumetric efficiency of the scavenging pump is constant regardless of the rotational speed of the engine, blowing-through of scavenging mixture will occur when the engine is operating at a rotational speed lower than the matching rotational speed, while exhaust gases will remain in the power cylinder when the engine is operating at a rotational speed higher than the matching rotational speed.
The volumetric efficiency of a reciprocating piston type scavenging pump means such as crankcase compression and a reciprocating type pump cylinder--piston assembly is generally higher as its rotational speed is lower; that is, when the pump means is directly driven by an engine, as the rotational speed of the engine is lower. Actually, however, an intake system which includes a carburetor connected to a scavenging pump means, an intake passage, and other flow resistant means such as reed valves incorporated in the intake passage, is subject to a pulsating effect; and, since such a pulsating effect changes in accordance with engine rotational speed, the volumetric efficiency of the scavenging pump means does not necessarily increase in accordance with decrease in engine rotational speed. Generally, in fact, the volumetric efficiency of a scavenging pump means is better when engine rotational speed is not very low but is relatively high. Further, when a larger number of pump cylinder--piston assemblies are operated with phase differences therebetween, the maximum volumetric efficiency is obtained at a higher rotational speed of the engine.
Generally, when an intake system operates with a single barrel type carburetor, the intake volumetric efficiency of a scavenging pump served by this caburetor changes in accordance with an upwardly convex curve as shown in FIG. 1, wherein the efficiency first increases in accordance with increase of engine rotational speed so as to reach the maximum value at a certain engine rotational speed, and then lowers when engine rotational speed further increases, provided that the venturi diameter is constant and the throttle opening is constant, as, for example, when the throttle is set at full throttle. In this case, the point of maximum volumetric efficiency shifts in the direction of high rotational speed as the venturi diameter is increased. However, increase of the venturi diameter means reduction of the speed of air flowing through the venturi, and causes poor atomization of gasoline in low load operation of the engine, thereby reducing engine torque, increasing fuel consumption, and reducing accelerating performance. Further, when the scavenging pump includes a plurality of chambers, the poor atomization of gasoline also causes poor distribution of fuel between these pump chambers and the power cylinders separately connected to these pump chambers. Therefore, when an engine is equipped with an intake system having a single barrel type carburetor and a venturi of a certain diameter so that the maximum intake volumetric efficiency of a scavenging pump driven by the engine is obtained at a medium point in the range of engine rotational speed, a difficulty is encountered with regard to matching of the power cylinder for scavenging. That is, if the power cylinder is matched for scavenging at an engine rotational speed which provides the maximum intake volumetric efficiency of the pump, as the rotational speed of the engine increases beyond this matching rotational speed, the scavenging becomes more and more incomplete. However, if the power cylinder is matched for scavenging at a high engine rotational speed, as the rotational speed of the engine lowers, more blowing through of scavenging mixture will occur.
On the other hand, in order to obtain good atomization of fuel over a wide range of engine rotational speed, it is already known to employ a two stage two barrel type carburetor which is opened or closed in two stages in accordance with engine load in such a manner that, when engine load is relatively low, intake air is drawn only through its first stage barrel, and when engine load increases its second stage barrel is also opened so that air is drawn through both the first and second stage barrels. In these conventional two stage two barrel type carburetors the first and second stage barrels generally incorporate first and second throttle valves, respectively, of which the first throttle valve which controls opening of the first stage barrel is directly operated by an accelerating mechanism including an accelerator pedal, while the second throttle valve which controls opening of the second stage barrel is interconnected with the first throttle valve by a link mechanism so that the second throttle valve is opened when the first throttle valve is opened beyond a predetermined opening, or alternatively the second throttle valve is opened by a diaphragm means which is operated by the venturi vacuum in the first stage barrel so that the second throttle valve is opened when the venturi vacuum in the first stage barrel increases beyond a predetermined value due to increase of air flow speed in the first stage barrel over a predetermined value. When scavenging mixture is supplied to a scavenging pump by employing this conventional two stage two barrel type carburetor, it is possible to set the engine rotational speed at which the intake volumetric efficiency of the pump becomes maximum at a point of high engine rotational speed, as shown by a broken line in FIG. 1, without deteriorating atomization of fuel in low speed operation of the engine, so that the matching of the power cylinder for scavenging is accomplished more desirably. However, the change of pump intake volumetric efficiency relative to engine rotational speed obtained by the conventional two stage two barrel type carburetor is still relatively moderate as seen in FIG. 1, and in this case, if the power cylinder is matched for scavenging at its maximum rotational speed, in low load operation considerable blowing-through of scavenging mixture still occurs.